This invention relates generally to swashplate type axial-piston hydraulic pumps, and in particular to innovations which increase the efficiency, adjustment range and speed capability and reduce the noise, size, weight and cost of such pumps.
Swashplate type axial-piston hydraulic pumps are well known in the art and typically include a generally cylindrical cylinder barrel rotatably mounted within a pump housing. One or more pump piston bores, having pump pistons reciprocably mounted therein, are disposed around the rotational axis of the cylinder barrel in parallel, or almost parallel alignment therewith. The ends of the pistons project beyond the end of the cylinder barrel so as to engage the surface of an angled swashplate stationarily mounted adjacent the end of the cylinder barrel within the pump housing. When the cylinder barrel is rotated within the housing, shoes, mounted to the piston ends, follow the surface of the angled swashplate with the result that the pistons are reciprocated within their respective piston bores. A valve plate, disposed adjacent the end of the cylinder barrel furthest from the swashplate, controls the ingress and egress of hydraulic fluid from the piston bores such, that a pumping effect is produced in response to rotation of the cylinder barrel within the pump housing.
Although highly advantageous in various applications, swashplate type axial-piston pumps are presently somewhat inefficient and their operational adjustment and speed range is too narrow when used e.g. as a vehicle transmission. (The adjustment range being the ratio of maximum to minimum swashplate angle which can be used efficiently). In addition, hydraulic pumps are generally too large, heavy and noisy at high power throughput and costly.
The inefficiencies are caused by friction due to high mechanical contact forces and leakage. These forces are representing mechanical loads like the side forces between piston and piston bore, retainer plate and shoe and retainer plate and retainer ring, or they are rest forces of loads which are hydrostatically balanced like the forces between shoe and swashplate, in the joint of the shoe and piston and cylinder barrel and valve plate. Furthermore, the friction force components of the mechanical contact forces produce tilting or cocking especially between the shoe and swashplate and the cylinder barrel and valve plate. Components of the piston side forces will increase the tilting between the cylinder barrel and the valve plate. These movements result in noticeable leakage and wear.
Attempts have been made to reduce the tilting or cocking of the cylinder barrel by applying counter forces which balance or nearly balance the tilting moment. Henry-Biabaud (U.S. Pat. No. 3,444,690) tries to balance axial forces and side forces of the floating spherical distributor/valve plate with radial forces of a reduction gear, located at the outer periphery of the cylinder barrel.
Further attempts have been made to reduce the tilting by supporting the cylinder barrel through a bearing located in the plane of the piston side forces on the shaft. The pulsating, resulting piston side force results in radial vibration of the shaft and therefore high frequency, small amplitude tilting of the cylinder barrel at the valve plate, creating additional leakage and wear.
Several prior attempts have been made to overcome the tilting and cocking of the shoe in relation to the face of the swashplate resulting in leakage and wear. The friction in the joint prevents the shoe from adapting fully to the face of the swashplate. A slight tilting is required, resulting in an excentricity of the mechanical force at the face of the shoe which overcomes the moment of friction in the joint These attempts to reduce the required moment and therefore the degree of tilting are directed toward the reduction of friction forces in the piston joint and increased countermoments through excentric hydraulic and mechanical forces at the face of the shoe and mechanical forces at the backface of the shoe due to a spring loaded or form locked retaining mechanism.
Previous attempts to reduce the moment of friction at the joint have lead to the increase of the pressure field within the joint to reduce the mechanical contact force by increasing the hydrostatic force, or to minimize the ball diameter with high friction forces on a small lever arm. Both solutions have had only limited success. The moment of friction on a small ball, limited by a sufficient encirclement of the socket around the ball, typically 20 or more past the geometric center of the ball, and the neck diameter between ball and piston, remains high, and a ball, noticeably larger than the piston diameter, providing space for a sufficient pressure field to reduce the mechanical contact forces, cannot be received deeper into the piston bore, therefore creating noticeable higher piston side forces or a reduced swashplate angle due to a longer lever arm between piston joint and piston bore.
Several prior art attempts have been made to create a sufficient excentric force at the face of the shoe to overcome the moment of friction of the piston joint to avoid or reduce the tilting of the shoe. This has been achieved through a not fully hydrostatically balanced axial shoe force creating a mechanical contact force, acting on the face diameter of a slightly tilted shoe, or through hydrostatic forces created through several pressure fields larger than needed which are partially depressurized due to the tilting of the shoe, creating an off-center hydraulic force. (Pat. SU 1421-894-A1). Various previous attempts have been made to provide these fields with a sufficient amount of fluid and pressure without producing an excessive amount of leakage, instability of the shoe movement, difficulties in fabricating the throttle arrangements and high sensitivity to wear or contamination. (Thoma, UK Pat. 983.310; Pat. SU 1463-951-A).
Both methods to create a counter moment to the friction moment at the joint do not produce the changing counter forces, needed at various swashplate angles and rotational positions during each revolution. Therefore, the remaining rest force or hydraulic forces are oversized at smaller swashplate angles and result in additional losses. The effects of the retaining mechanism are stated later.
Several attempts have been made to reduce the leakage and wear sensitivity at the shoe. Deflecting ends at the face of the shoe have been utilized to provide a hydrodynamic pressure field, thus reducing the size and leakage of the required hydrostatic pressure field. (Espig et al, U.S. Pat. No. 3,521,532). High strength materials for the shoe, such as steel, have been utilized with enclosed bearing material at the face of the shoe to reduce deterioration during service. (Alexanderson et al, U.S. Pat. No. 3,263,623). In both attempts, the shoe socket end is fitted over the ball at the piston by 230 or more.
Prior attempts to reduce the frictional losses between piston and piston bore have been directed toward establishing hydrodynamic or hydrostatic pressure fields at the contact areas. The variants have typically been the clearance between piston and piston bore and the elasticity and/or shape of the ends of the piston bores to improve the conditions for a hydrodynamic pressure field or to establish a hydrostatic pressure field. (Thoma, U.S. Pat. No. 3,216,333).
Additional losses occur due to mechanical preloads (spring force) between the cylinder and the valve plate and the retaining mechanism, shoe and swashplate. The spring load is needed to hold the parts in position at no or very low pressure rates and to provide additional forces at the back face of the shoe to reduce its tilting. The pre-load forces are generally constant and result in high percentile losses at low pressure rates and small swashplate angles.
Several attempts have been made to use form-locked retaining mechanisms for the shoes to eliminate the effects of the pre-load forces in several sections of the mechanism, between shoe and swashplate, shoe and retainer plate and retainer plate and its joint mechanism consisting generally of a spring-loaded ball.
Prior art designs of form-locked retaining mechanisms surrounding the drive shaft are located at the outer circumference of the mechanism or do not allow the shaft to be extended through the swashplate. They are space consuming, do not provide sufficient stiffness to hold the shoes in their desired position and result in higher bearing losses and costs.
Another attempt has been made to reduce the tilting of the shoe by coupling the shoe to a sliding disk, running on the face of the swashplate. (Riedbammer, U.S. Pat. No. 5,056,403). The sliding disk subassembly, containing all elements of a form-locked mechanism, is pressed with spring force against the swashplate. This form-locked subassembly mechanism reduces the tilting of the shoe relative to the sliding disk, but increases the number of moving parts, high-pressure sealing areas, sensitivity to contamination, cost and friction and space requirements due to the spring force.
Another major loss occurs during the pressurization and depressurization of the pumped fluid and the gases which are contained in the piston bore and the bore channel. Prior art axial-piston pumps have swashplates rotating about a centrally located axis. This results in piston strokes which are symmetrical about their zero degree swashplate angle position. This produces an increasing unswept piston bore volume with decreasing swashplate angle. Therefore, the compression losses are most critical at smaller swashplate angles and higher pressure rates when the ratio of compressed fluid to pumped fluid is high. This contributes significantly to the inefficiency and low suction capacity of an axial piston pump. Some previous attempts have been made to reduce these compression losses by beginning the suction stroke always at the top dead-center position of the piston. Typically, the solutions for all types of axial-piston pumps have been relatively expensive, space consuming, and heavy (Ifield, U.S. Pat. No. 4,129,063), are mechanically not reliable (Bosch, U.S. Pat. No. 3,733,970), do not allow the reversal of the flow direction or not even a full adjustment between maximal and zero degree adjustment angle.
The present speed ranges are limited because of cavitation, occurring between valve plate and cylinder barrel due to high velocities and unfavorable flow patterns at high speeds. The minimum speed is determined by a decreasing efficiency and an increasing torque fluctuation. In typically prior art pumps of medium size, values above 3500 rpm and below 500 rpm are not considered to be practical. Thus the typical speed range of previous axial-piston pumps swashplate type, medium size is approximately 7 to 1.
The present adjustment range of an axial-piston pump swashplate type is limited because of excessive side forces and deflection of the piston in its most extended position in bottom dead-center. The minimum swashplate angle is determined by a decreasing efficiency. In typical prior art pumps, the maximum swashplate angle is 15 to 20, typically 18, and the minimal angle is approximately 7 to 8. Thus the typical adjustment range is approximately 2.5 to 1.
Several attempts have been made to reduce the extension of the piston from the face of the cylinder barrel in bottom dead center. (Friedrich et al, Germany/BRD OL 1954565; Takai, U.S. Pat. No. 4,776,259). Both provide a circumferential relief at the piston bore end at the side of the swashplate, thus providing a deeper reception of the piston and its joint into the piston bore. The reduction of the effects of the piston side forces is marginal since the relief is circumferential, providing very limited or no additional support for the piston.
The development of noise in pumps or motors results from abrupt changes of forces due to abrupt pressure changes in the piston bore when rotating from one valve plate port to another. Prior art designs have basically attempted to delay the pressure change by providing grooves in circumferential direction as extension of the ports. These grooves are noticeably effective only at certain points of operation, varying because of different swashplate angles, speeds, fluid viscosities and pressure ranges. In addition, the grooves increase the internal leakage and therefore reduce the efficiency.
The size and weight of axial-piston pumps and motors of prior art design are too high to be used economically as transmission component in automotive applications, especially when used as a motor. Presently, typical adjustable axial-piston pumps have a power to weight ratio of approximately 2.5 to 1 (hp/lbs.).
It is therefore desirable to increase the efficiency and the transformation ratio (adjustment and speed range) and to reduce the noise, size, weight and cost of an axial-piston pump by overcoming these and other problems in the prior art.
An improved swashplate type axial-piston pump has increased efficiency, a greater transformation ratio (adjustment and speed range), is smaller in size and weight, develops less noise and is less costly to make it suitable for a wider range of applications, especially for the use as an automotive transmission.
In a preferred embodiment, the piston assembly includes a spherical joint. Socket and ball of this joint are machined to their final shape before they are meshed together. Thus the need to deform one or both parts during the assembly process is eliminated and high strength material can be utilized. This snap-fit joint results in a larger joint with reduced mechanical contact forces and an improved contact surface for less friction, less leakage and reduced cost.
In a preferred embodiment, the piston joint assembly includes a throttle means for balancing or reducing the mechanical axial forces between shoe and swashplate and within the joint between shoe and piston. The throttle means preferably includes a first conduit means in the piston for transferring hydraulic fluid from the piston bore to a first end of the piston, and a second conduit means in the shoe for transferring the fluid from a first shoe end to the swashplate upper surface. A channel means is also provided at the first piston end and the corresponding first shoe end surface for transferring the hydraulic fluid. The channel means at this piston joint surface may have one of several configurations. It may include one or several concentric grooves in one or both, the piston or shoe end surface which may be connected by a passage. Instead, it may include a helical shaped groove in the surface of either the piston or shoe end surface with a concentric groove at the opposite surface. The channel means results first, in an increased high pressure field at the joint, reducing the mechanical contact force and therefore the moment of the joint friction, and second, a hydrostatic pressure field between shoe and swashplate supplied with varying, continuously or intermittently changing pressure rates (considering a comparable flow of leakage at the contact area) proportionally or nearly proportionally with the varying axial shoe force, thus minimizing the remaining mechanical contact force and the friction between shoe and swashplate and the leakage at all swashplate angles.
Due to reduced side forces at the piston in its most extended position, the preferred embodiment includes piston bores having notches in radial direction near the ends of the bores at the side of the swashplate The notches allow the joint and the neck of the shoe to be received deeper into the piston bore at the top dead-center position. This arrangement reduces the contact forces between the piston and piston bore because of a reduced lever arm between piston joint and the onset of the piston bore in bottom dead-center and no tilting forces at the piston after the joint has entered the piston bore. This arrangement allows a larger swashplate angle. The effect of this arrangement at smaller swashplate angles is even greater when used in combination with the off-center swashplate adjustment means discussed later.
The undesirable pre-load forces of the retaining mechanism of prior art designs are minimized in the present invention by providing a form-locked retainer means which retains the shoe in its desired position against the swashplate upper surface. In a preferred embodiment, the retainer means includes a retainer ring or collar that substantially surrounds the pump shaft, and a retainer plate that engages both the retainer ring and shoe. The provision of an internal retainer ring near the shaft increases the amount of usable space at the outer periphery of the swashplate, especially when utilizing a spherical face at the swashplate, permitting a larger swashplate angle, increased stiffness of the mechanism and reduced frictional losses.
In a preferred embodiment, the retainer plate has a substantially spherical upper surface to match the opposite surface at the retainer ring. Furthermore, if a swashplate with a spherical surface is utilized, all mating spherical faces at the swashplate, the shoes, the retainer and the retainer ring have substantially the same center point. This arrangement allows the retainer plate first, to be rotated about the shaft, following the rotational movement of the shoe, and second, to move normal to the shoe axis or swivel about the center point of the spherical surfaces, following the centerlines of the shoes, resulting in a tilt angle between the centerline of the retainer plate and the cylinder barrel that is larger than the swashplate angle. This retaining means allows a smaller bore for the shoe neck in the retainer plate, resulting in improved guidance for the shoe, an increased swashplate angle due to reduced space requirements of the retainer plate in radial and axial direction and when combined with a smaller pump shaft diameter, sufficient space for an internal retainer ring.
The pump according to the present invention has a high speed capacity because of an increased size of the piston bore channel, tilted inward and in circumferential direction, therefore reducing the flow velocity and the turbulence. This is accomplished by a reduced pitch diameter of the valve plate ports and the corresponding bore channel openings.
The area of the valve plate port containing high pressure and the bore channel openings connected with the port create a pressure field whose centroid is distanced from the centroid of the combined hydraulic forces of the piston bores, or reaction forces of the axial piston forces, connected with the port, therefore creating a tilting moment at the cylinder barrel. This tilting moment is substantially compensated by a counter rotating tilting moment created by the combined radial force at the piston joints acting perpendicular to the plane of the centerlines of the pistons in dead-center positions and its distance to the equivalent force point of the cylinder barrel bearing.
To improve the efficiency and reduce the noise, the valve plate has two compensating ports in fluid connection with each other to transfer part of the decompression volume from the high pressure piston in its top dead-center position to the low pressure piston in its bottom dead-center position. This reduces the compression and decompression losses of the pistons in top and bottom dead-center position, their forces when they do not produce a noticeable amount of torque at the shaft (as motor) or fluid flow (as pump) and reduce the development of noise due to a stepwise decrease or increase of fluid pressure, especially when utilizing an even number of pistons for the cylinder barrel.
The present invention includes an off-center, dual axis adjustment mechanism for the swashplate that tilts around an axis, located near the centerline of the piston in top dead-center position. There is an axis for each tilting or flow direction, represented by swivel mechanism with two joints, connecting the swashplate to adjustment plungers. Due to stops at the swashplate and the plungers, the swashplate rotates, starting in neutral or zero degree position, about the plunger joint of the swivel mechanism and then about its swashplate joint. Thus, the center of the swashplate face, starting at the centerline of the shaft, describing two arcs during a complete tilting movement, remains close to the centerline of the shaft.
This tilting movement results in a piston stroke which begins always at the maximum of the top dead-center position and provides minimized dimensions for the retainer ring and retainer plate.
In addition, the plunger provides support for forces of the swashplate in radial direction of its centerline created through side forces of the piston assemblies acting on a spherical face of the swashplate and support against rotation, resulting from the friction between the shoe and the swashplate. A minimum of three joint links on two axes is provided, holding the resulting piston forces of the high and the low pressure section at or within the frame of their support joint This prevents an undesirable cocking of the swashplate around the plane of the centerlines of the pistons in top and bottom dead-center. Another advantage of this arrangement is, that only one swashplate axis is moving while the other remains in its zero-position, simplifying the control of the swashplate adjustment.
It is another feature and advantage of the invention to reduce the weight, size and cost of an axial-piston pump.
These and other features and advantages of the present invention will be apparent to those skilled in the art from the following detailed description of the preferred embodiments and the drawings in which: